1. Field of the Invention
The invention is directed to a flat plate air-to-air heat exchanger having a heat exchanger core formed by using gasket strips and interlocking clips to structurally join enhanced heat transfer surfaces. Uniquely shaped thin flat plate metal sheets create flow channels and heat transfer surfaces necessary to promote a combination of cross-flow and counter-flow heat exchange between two air streams to yield both a thermally efficient and cost-effective heat exchange configuration.
2. Description of Related Art
Heat exchangers, of which air-to-air heat exchangers are a subset, have been widely used in industrial processes, energy systems, and air conditioning and refrigeration systems for many years where there is a need to transfer energy between two fluid streams to achieve the system or process objectives without having direct contact between these two fluid streams. Heat exchangers using metal cores are capable of operating at elevated temperatures and pressures. These heat exchangers also exhibit high efficiency heat transfer. Flow and pressure boundary joints of metal heat exchangers are typically welded or formed by bending the metal to form a mechanical joint that is allowed to leak or is sealed with joint compound. It is widely known that the counter-flow (fluids flow in opposite directions) heat exchanger configuration is more efficient (requires the least amount of surface area to transfer the heat load) than the cross-flow configuration (fluids flow perpendicular to each other). Since a square or rectangular plate has four (4) sides or two pair perpendicular to each other, it is straight forward to configure a crossflow heat exchanger and much more complex to configure a counter flow heat exchanger to convectively exchange heat between two air streams.
ASHRAE Handbook, Chapter 43 ‘HVA Systems and Equipment’, provides a summary of the current art of heat exchanger design and fabrication. In particular, this reference points to shell and tube, double pipe, finned tube, spiral plate, plate and frame, and plastic heat exchangers as the current configurations of heat exchangers covered by the current art. ASHRAE Handbook, Chapter 4 ‘Air-to-Air Energy Recovery’, provides a summary of the state of the art heat exchange devices deployed to recover energy from an airstream at a high temperature to an airstream at a lower temperature. Included in the discussion of these devices is the air-to-air fixed plate heat exchanger which is available in many configurations, materials, sizes, and flow patterns. The plates are formed with spacers or separators (e.g., ribs, dimples, ovals) constructed into the plates or with external separators (e.g., supports, braces, corrugations). Airstream separations are sealed by folding, multiple folding, gluing, cementing, welding, or any combination of these, depending on the application and manufacturer.
The plates are thin such that the heat transfer resistance through the plates is small relative to that of the airstream boundary layer on each side of the plate. Therefore, the material selected for the heat exchange surface does not substantially impact the heat transfer efficiency. Plate spacing ranges from 0.1 to 0.5 inches depending on design and application. The cross-flow heat exchangers are used most often because of their lower cost even though counterflow heat exchangers increase heat transfer and energy efficiency.
The general literature shows the air-to-air heat exchanger to be a reliable low pressure passive heat transfer device with little or no direct communication between the air streams. Additionally, the obvious point is made by these publications that flat metal sheet stock costs less per unit of surface area than tube stock and that this flat sheet stock offers many possibilities for fabricating heat exchanger flow passages that are aerodynamically and hydraulically effective to support fluid flow while minimizing pressure loss. This literature also shows the limiting weaknesses in using flat sheet stock to be: (a) the stress concentration at joints, (b) exchanger size, (c) flow distribution, (d) limited capacity to handle pressure differentials and (d) cleaning of heat transfer surfaces.
The general physical behavior of fluid flow and heat transfer as well as the specific behavior for many configurations has been characterized previously through experiments and observations. Pressure drop and heat transfer are critical performance characteristics of heat exchangers. Both of these performance characteristics are strong functions of the velocity and profile of the airflow as it enters, traverses, and exits the heat exchanger as driven by the physical arrangement and materials of the heat exchanger. Reynolds number and Nusself number are commonly used to quantify local fluid flow and heat transfer characteristics for further characterization of the overall thermal and physical performance of heat exchangers. Reynolds number is a dimensionless group of the ratio of inertial force (free stream velocity) to fiscous force (boundary layer force). Nusself number is the dimensionless ratio found by dividing the convective conductance or heat transfer coefficient, h, by the molecular thermal conductance, k/L, for flow over a surface, or k/de, for flow in a channel or duct with k representing the thermal conductivity of the fluid, L representing flat surface length and de representing the hydraulic diameter of the channel, for the configuration the heat transfer system. The Darcy-Weisbach equation factors in the Reynolds number combining it with relative roughness to determine the duct friction factor to characterize the heat exchanger pressure loss. Namely,Δp=fD(L/d)(rv2/2)                Where:        Δp is pressure loss along the round pipe or duct        fD is the friction factor which is a function of Reynolds number and relative roughness        L/d is the geometric factor for length per unit diameter        r is the gas density        v2 is the velocity squared        
Heat transfer in a heat exchanger is characterized by the general equationQ=UALmtd                 Where:        Q is the quantity of heat being transferred BTU/hr        U is the overall heat transfer coefficient which is dependent on conductance through the two air boundary layers at the plate interface and the conductance through the plate BTU/(hr° F. ft2).        Lmtd is the log mean temperature difference between the two air streamsAnd U=1/(Ro+1/ho+t/k+1/hi+Ri)        where:        U=overall heat transfer coefficient        ho=hot-side film coefficient (BTU/(hr° F. ft2))        hi=cold-side film coefficient (BTU/(hr° F. ft2))        t=plate thickness (ft)        k=material thermal conductivity (BTU/(hr° F. ft)        Ro=hot-side fouling resistance ((hr° F. ft2)/BTU)        Ri=cold-side fouling resistance ((hr° F. ft2)/BTU)        
It is well known that the limiting factor in heat transfer in heat exchange involving an air side surface is the thermal resistance of the thin layer of low velocity air that hugs the heat transfer surface. Many methods have been employed to disturb this thermal boundary layer by making it thinner or by breaking it up into unique airflow streams. Various mechanisms have been deployed in today's advanced systems with the objective of accomplishing this and improving heat transfer including: pin fins, protruding ribs, louvered fins, offset-strip fins, slit fins, dimples, and others. These wall layer perturbations are meant to not only periodically break up this boundary layer but to agitate the wall layer such that there is improved mixing with the main air flow stream. A penalty of these heat transfer augmentation mechanisms is the increase in hydrodynamic resistance which requires more power to propel the air through the heat transfer channel. It has been shown for example that louvered fins increase the heat transfer by a factor of between two (2) to five (5) but the friction losses increase by a factor of between four (4) to ten (10). Since fan power increases as the square of increased pressure drop such an increase in friction loss is a severe penalty for increased heat transfer effectiveness.
Thermal efficiency of heat transfer augmentation, as characterized by the ratio of ending to beginning heat transfer (Nusself Number) divided by the one third power of ending to beginning friction factor (((Nu/Nuo)/(f/fo)1/3), shows dimples to be one of the more effective heat transfer augmentation method (Dimples=2.2, Louvered finx=1.6, Pin fins=1.5, and Ribs=1.0)a. Since 2005 the research focus, as documented by a number of researchers, has been on using concavities and dimples, to provide increased heat transfer. The bulk of this research has been conducted through a combination of experiments using test fixtures with single sided channels and computational numerical methods. Heat transfer enhancement tests at the National Research Laboratory Program of KISTEP (Korea Institute of Science and Technology Evaluation and Planning) were conducted using three different roughened surfaces of dimpled, protruded, and complex (dimple-protrusion) in a rectangular channel with a width to height aspect ratio (W/H) of 7. The diameter, depth, and pitch of both the protrusion and dimple is the test rig were selected to match the findings of previous research which identified the best performing values for these variables. The test results show: (a) the high heat transfer region at the rear side of the dimple caused by increased flow mixing induced by upwash flow and pairs of vortices with recirculation flow inside the dimpled surface creating a low heat transfer region; (b) heat transfer enhanced by the impinging effect of the horseshoe vortices on the front side of the protrusion; and (c) compound heat transfer characteristics of impingement and upwash vortices occur for the complex dimple and protrusion configuration. These tests show the lowest friction factor for the dimpled case with the complex case being 3 times higher, and the protrusion case being 6 times higher. A conclusion of the test is that when the dimple/protrusion are installed at both walls, the cross-sectional area of the channel decreases, especially for the protruded case, and that this leads to increased system pressure loss.
U.S. Pat. No. 6,076,598 describes an opposed flow (counter flow) heat exchanger that uses corrugated and flat plate elements arranged to form heat transfer surface and air flow passages. The materials of construction are paper, wood pulp, organic material and metallic foil. The inlet and exit passages (called out in FIG. 1 as #7 and #8) are located at the ends of the heat exchanger and are formed by the heat exchanger skin and the flat plates that run the full length through the heat exchanger core. The corrugated plates are placed on top of and joined to the flat plates with the corrugated plates running the full width but not the full length of the flat plate. The gap between the end of the flat plate and the end of the corrugated plate acts as an air flow entrance or exit passage and distribution/collection header. The corrugated plates form the primary heat transfer surface.
Air has to make many changes in direction as it moves from entrance to exit resulting in large pressure loss (entry, distribution, core area flow, collection, and exit). This flow path limits the flow rates through the heat exchanger leading to the high likelihood of areas with laminar flow and poor heat transfer. Poor heat transfer results in the need for more area and therefore a larger heat exchanger.
Joining each of the corrugated sheets to the flat plates and further shaping and joining the flat plates to form the entrance and exit pathways and the flow channels adds significant complexity and cost to the manufacture of this heat exchanger.
U.S. Pat. Nos. 5,072,790 and 4,848,450, by the same inventor describe a cross-flow heat exchanger where the heat exchanger is made-up of multiple rectangular plates. These plates are formed by turning flanges on each side that establish a specific and uniform standoff height or gap between plates. The plates interlock by sliding the plate flanges together to create an air flow channel and the heat transfer surface. Recognizing that the mechanical process for forming the flanges on large heat exchangers introduces a variation in the space between the plates and a twisting of the plates as they are assembled, the invention was upgraded in U.S. Pat. No. 5,072,790 with the introduction of side and corner support members. These members are designed to more effectively control the gaps between the plates from the edge of the heat exchanger.
The efficiency of this heat exchanger is limited by its design as a cross-flow heat exchanger which limits the heat exchanger's effectiveness/efficiency. Additionally, the side and corner spacer support members are located on the perimeter of the heat exchanger such that the gap between the flat plates can vary as you move to the center of the heat exchanger core. This variation impacts the flow distribution and therefore the heat transfer rate across the plate.
U.S. Pat. No. 4,554,719 describes a cross-flow heat exchanger and the method for manufacturing this heat exchanger. The heat exchanger is formed by a stack of aluminum plates with flanges formed to create lap-joints which hold the plates together and create separation between the plates to form the flow channel for the heat exchanger. Nodules are formed in the aluminum plates to improve air turbulence and heat transfer. The invention centers on the method for forming a heat exchanger core that has lap joints and nodules on rectangular plates with tooling designed to form these shapes in a single step.
The efficiency of this heat exchanger is limited by its design as a cross-flow heat exchanger. The nodule design characteristics with associated impact on pressure drop and Reynolds number as a function of mass flow rate is not accounted for in the method or the design.
U.S. Pat. No. 4,350,201 describes a cross-flow plate and fin heat exchanger having features that act of self fixture the unit for brazing. The unit is composed of conventional elements conventionally stacked one upon another in a standard cross-flow configuration. A corrugated plate is shown as the fin element. The flat plate elements are uniquely configured to constrain and position the fin and spacer elements to achieve the desired configuration to enable brazing or welding without additional positioning or anchoring.
The efficiency of this heat exchanger is limited by its design as a cross-flow heat exchanger. The corrugated plate acting as a fin surface is sandwiched in between two flat plates with no structural or conduction bonding between the corrugated plate and the flat plates. This lack of bond to the flat plates limits the effectiveness of the fin surface. The entrance an exit losses are high with the contraction and change of direction.
U.S. Pat. No. 5,785,117 describes a cross-flow air-to-air heat exchanger core composed of square plates with flanges on each end of the plate where the two opposite sides are bent in the same direction and opposite the direction of the adjacent flanges, the flanges being formed by bending the ends of the plates to 90 degrees.
The efficiency of this heat exchanger is limited by its design as a cross-flow heat exchanger. No provisions have been made to address the inlet and outlet pressure losses. No provisions have been made to reduce air boundary layer heat flow resistance for laminar or transition conditions.
U.S. Pat. No. 4,125,153 describes a cross-flow plate air-to-air heat exchanger core composed of square plates with flanges on two sides of the plates formed by bending the ends of the plates to 90 degrees.
The efficiency of this heat exchanger is limited by its design as a cross-flow heat exchanger. No provisions have been made to address the inlet and outlet pressure losses. No provisions have been made to reduce air boundary layer heat flow resistance for laminar or transition conditions.
The synopsis of the state of the art, recent research, and selected most relevant patents above show (a) past manufacturing limitations have generally limited flat plate heat exchangers to be crossflow heat exchangers; (b) the materials and methods used to join the heat transfer plates and form the pressure boundary have been limited to formed or welded mechanical joints; (c) heat transfer surface augmentation is possible and beneficial; (d) heat transfer augmentation approaches to date have included louvered fins, pins, protrusions, and dimples; (e) augmentation approaches to date have increased heat transfer with a significant pressure drop penalty; and (f) the size, weight, and cost penalties for air-to-air heat exchange remains significant.